1. Field of the Invention
The present invention relates to the conversion of chemical energy (fuel) into hydraulic, electric or pneumatic energy. The general field of application is the efficient production of hydraulic, electric or pneumatic power for mobile and non-mobile power needs.
2. The Prior Art
Hydraulic power is currently produced by rotating the drive shaft of a hydraulic pump by a drive motor, usually an electric motor or an internal combustion engine. Power from a rotating shaft must be converted into a linear motion to drive reciprocating pistons which serve as the pumping means for the most efficient hydraulic pumps. When a reciprocating piston pump is driven by a conventional crankshaft internal combustion engine, pistons within the engine are driven linearly by the expansion of combustion gases, which in turn are connected by rods to a crankshaft, to produce rotating power output, which in turn is connected to the drive shaft of a piston pump which must then create the linear motion of the pumping pistons to produce hydraulic power.
The idea of directly (and usually axially) coupling the engine combustion piston to the hydraulic piston to produce hydraulic power directly from the linear motion of the combustion piston, avoiding the cost and inefficiencies of converting linear motion to rotation and back to linear, is not new. However, a variety of challenges associated with prior art designs have prevented any commercial success of this basic idea.
Connecting the combustion piston to the hydraulic piston eliminates the need for an engine crankshaft, and in doing so forms a free-piston assembly. Since the piston assembly is not connected mechanically to an apparatus which could in turn be used to control thernovement of the free-piston assembly, one major challenge associated with the basic idea of free-piston engines is how to accurately and repeatably (for millions of events) control the exact position of the stoppage of the assembly as it approaches the top dead center (TDC) position of the combustion piston during its compression stroke. For a combustion engine to be efficient, the control of the degree of compression (that is the compression ratio) is critical, and the high compression ratios of efficient combustion processes result in the need to take and stop the combustion piston very near (often within 1 millimeter) the opposite end of the combustion chamber (usually the engine xe2x80x9cheadxe2x80x9d). A similar challenge is associated with the control of the exact position of the stoppage of the assembly as it approaches the bottom dead center (BDC) position of the pumping piston during the expansion or power stroke. The friction of each stroke can vary (especially during warm-up or transient operation), the quantity of fuel provided for each combustion event can vary, the beginning of the combustion process can vary, the rate of combustion and its completeness can vary, the pressure of the hydraulic fluid being supplied to the pump can vary, the pressure of the hydraulic fluid being expelled can vary, and many other operating parameters that influence each stroke can vary; therefore, the accurate control of the TDC and BDC positions is very challenging. The consequences of inadequate control can go beyond unacceptable performance, and be destructive to the engine if the combustion piston contacts the opposite end of the combustion chamber or the pumping piston contacts the opposite end of the pumping chamber.
Free-piston engines of the prior art operate on the two stroke cycle (with one exception to be described later) because of the challenge of operational control. Even with a two stroke cycle, stoppage of the combustion piston at the correct position at TDC during the compression stroke is very difficult. If the engine were operating on the four stroke cycle, an additional TDC stroke would be required to exhaust the spent combustion gases. In this exhaust stroke, unlike the compression stroke, there would be no trapped gases to increase in pressure as the combustion piston moved toward TDC and thereby decelerate the piston assembly. Some other means would be necessary to restrain the piston assembly from impact. Additional means would also be needed to move the assembly through the two additional strokes. Other problems or disadvantages of prior art designs will be apparent as they are contrasted with the present invention.
There are several informative technical papers, Society of Automotive Engineers (SAE) papers numbers 921740, 941776, 960032 and the reference listed therein, which provide review and analysis of the various free-piston engine concepts. There are also several United States free-piston hydraulic pump and related technology patents which might be considered relevant to the present invention and are as follows:
U.S. Pat. No. 4,087,205 Heintz: Free-Piston Engine-Pump Unit
U.S. Pat. No. 4,369,021 Heintz: Free-Piston Engine Pump
U.S. Pat. No. 4,410,304 Bergloff et al: Free Piston Pump
U.S. Pat. No. 4,435,133 Meulendyk: Free Piston Engine Pump with Energy Rate Smoothing
U.S. Pat. No. 3,841,707 Fitzgerald: Power Units
U.S. Pat. No. 6,152,091 Bailey et al: Method of Operating a Free Piston Internal Combustion Engine
U.S. Pat. No. 5,983,638 Achten et al: Hydraulic Switching Valve, and a Free Piston Engine Provided Therewith
U.S. Pat. No. 5,829,393 Achten et al: Free Piston Engine
U.S. Pat. No. 4,891,941 Heintz: Free-Piston Engine-Pump Propulsion System
U.S. Pat. No. 4,791,786 Stuyvenberg: Free-Piston Motor with Hydraulic or Pneumatic Energy Transmission
U.S. Pat. No. 4,382,748 Vanderlaan: Opposed Piston Type Free Piston Engine Pump Unit
U.S. Pat. No. 6,029,616 Mayne et al: Free Piston Engine
U.S. Pat. No. 5,556,262 Achten et al: Free Piston Engine Having a Fluid Energy Unit
U.S. Pat. No. 5,363,651 Knight: Free Piston Internal Combustion Engine
U.S. Pat. No. 5,261,797 Christenson: Internal Combustion Engine/Fluid Pump Combination
U.S. Pat. No. 4,415,313 Bouthors et al: Hydraulic Generator with Free Piston Engine
There is also a free-piston, hydraulic-pump engine, which can operate in either the two stroke or four stroke cycles, disclosed in U.S. Pat. No. 5,611,300 (FIGS. 6-8 and claims 11-12). This engine utilizes a conventional crankshaft and combustion piston to intake and compress air and to exhaust the spent combustion gases for the four stroke cycle.
Free-piston engines of prior art design are generally classified as single piston, opposed piston or dual piston. The present invention would be classified as a dual piston configuration. Like prior art free-piston engines, the present invention utilizes the stroke of the combustion piston to directly produce hydraulic, pneumatic or electric energy. However, for ease of description of the essential features of the present invention, only hydraulic energy production will be described.
Additional challenges associated with the various prior art free-piston engine designs include:
(1) Difficulty in achieving mechanical balance. Each stroke of a free-piston assembly transmits an acceleration and a deceleration force to the engine housing, and to the structure to which the engine is mounted unless these forces are somehow counteracted (i.e., balanced) within the engine. Proponents of opposed piston engines usually stress as a primary advantage the potential for good balance, but the difficulty of exactly controlling the movement of each free-piston makes this potential difficult to realize in practice.
(2) Accurate control of timing and quantity of fuel introduction. This challenge is primarily related to control of the piston assembly motion as previously discussed, but the elimination of this sensitivity would be highly beneficial.
(3) Operation utilizing two stroke cycle. There are currently no two stroke cycle automotive engines sold in the United States. This is because it is extremely difficult to control air pollution exhaust emissions from such engines. This challenge would apply to two stroke cycle free-piston engines as well.
(4) Difficulty of providing a wide range of power output. A natural frequency (similar to a mass-spring-damper system) exists for any type of free-piston engine, and it is difficult to significantly vary this speed. This natural frequency is influenced most by the mass of the piston assembly and the stroke length. Smaller values for mass and stroke increase the frequency but greatly increases the velocity especially during the early part of the expansion or power stroke. The increased velocity in this region inhibits complete combustion and reduces the hydraulic efficiency of the pumping piston. In an attempt to increase frequency and thereby specific power, most prior art free-piston engines strive to minimize mass and thus incur combustion and efficiency penalties. To vary power output they teach intermittent operation. Operation can pause after each cycle so varying the pause time will vary the average power output. However, the time for each cycle was fixed by the high natural frequency, and the engine continues to experience the efficiency penalties previously mentioned.
(5) Difficulty of responding to varying high pressure levels. Most hydraulic systems where free-piston engines would be attractive experience a wide range in system high pressure levels, e.g., from 2000 to 5000 psi. Many free-piston engine designs would operate with a fixed high pressure and thus have limited applicability. Others would require changing the fuel supply level to correspond to changing pressures. For example, at 5000 psi the engine fuel consumption level (per cycle) would be maximum and proportionally lower at lower pressures. One obvious problem with this approach is that the hydraulic power output drops with pressure, e.g. at 2500 psi only one half the maximum power output could be supplied. Also, there is usually a need for increased (not decreased) power if the system pressure drops. Others have suggested using a well known pumping flow xe2x80x9cBypass systemxe2x80x9d (Beachley and Fronczak in SAE paper 921740) or by another name xe2x80x9ccoupling a hydraulic accumulator with said pressure chamber at a selected point in time during said return stroke to thereby attain said output operating pressurexe2x80x9d (U.S. Pat. No. 6,152,091) or by another name xe2x80x9cadjustment of the effective piston strokexe2x80x9d (U.S. Pat. No. 6,814,405, Octrooiraad Nederland). The size of the hydraulic pumping chamber is such that even at the lowest expected pressure (e.g., 2000 psi), the maximum combustion energy can be delivered as hydraulic flow through no more than the full stroke of the pumping piston. At higher pressures, a valve would bypass the initial flow back to the low pressure system, shutting that valve at a position in the power stroke where the remaining stroke is needed to transfer the full combustion energy to the high pressure hydraulic system. Theoretically, this approach would allow the engine to run at an optimum condition independent of system high pressure level. The bypass flow system has been used in several commercial, non free-piston engine hydraulic systems such as diesel engine fuel injection pumps and certain variable displacement xe2x80x9ccheck valvexe2x80x9d hydraulic pumps (e.g., Dynex pumps). For example, in diesel engine fuel injection pumps, a piston chamber is charged (much like the method of the piston chamber of free-piston engines), through a check valve with low pressure diesel oil from the fuel tank, as the piston moves from TDC to BDC within the piston chamber. Then, as the piston returns from BDC toward TDC, a xe2x80x9cspill valvexe2x80x9d allows fuel to bypass the high-pressure check valve outlet to the injector and return to the tank. Depending on the torque command (i.e., the fuel quantity needed for injection), the bypass valve will shut at the appropriate stroke position to deliver the needed fuel through the high pressure check valve to the injector. The reason that this approach to xe2x80x9cvarying the effective stroke of the pumping pistonxe2x80x9d has not yet been commercially successful in free-piston engines is because it results in an unacceptable efficiency loss. For the free-piston engine, the bypass flow rate is the highest flow rate in the cycle. This is because there is little resistance to the flow and the velocity of the piston is at maximum since the expansion of the combustion gases has accelerated the reciprocating mass to its maximum speed. After the bypass is shut, the pumping work decelerates the assembly. To provide xe2x80x9clittle resistancexe2x80x9d to this high flow rate, the bypass valve must be very large. If the valve is too small, the flow pressure losses will waste potential hydraulic power and greatly reduce efficiency. A large bypass valve on the other hand has a larger relative mass and, for a given closing force, will shut much slower. During the closing period the high flow rate experiences an increasing pressure drop and wastes potential hydraulic power. Existing systems utilizing this approach experience such losses. For the diesel engine fuel injection example, the power associated with the flow rate of the diesel fuel is so low relative to the power output of the diesel engine (or relative to the power associated with the flow rate for a comparable power free-piston engine) that some losses in efficiency have a relatively small impact on the diesel engine efficiency, although still significant and the subject of much research. Likewise, variable displacement check-valve hydraulic pumps are significantly less efficient than other approaches to varying displacement in hydraulic pumps, but because of their simplicity and relatively low cost, they have found some commercial success. For a free-piston engine to be successful in utilizing a bypass valve approach, it must operate with minimal open flow losses, be able to accurately and repeatably shut on command, and most importantly, must be extremely fast.
Prior art dual piston configurations of free-piston engines contain a pair of opposed power pistons which are fixedly, internally interconnected. Each power (combustion) piston has a hydraulic pumping piston axially attached through a connecting rod. FIG. 1 shows the free-piston assembly of prior art dual piston configurations. Opposed combustion pistons 2 and 3 slide within combustion cylinders (not shown). Combustion pistons 2 and 3 respectively have inwardly attached pumping pistons 4 and 5 which slide within pumping cylinders 6 and 7. The pumping pistons 4 and 5 are fixedly and internally connected through sealing block 8 by connecting rod 9, whereby combustion pistons 2 and 3 and pumping pistons 4 and 5 and connecting rod 9 reciprocate as a unit. Coaxially and therefore internally connecting a pair of single unit free-piston assemblies to form a dual piston assembly presents several problems:
(1) The free-piston assembly is longer than would otherwise be necessary by the length of sealing block 8.
(2) A high pressure hydraulic fluid seal (or pair of seals) must be provided within the sealing block 8 which adds cost and imposes increased friction which significantly reduces overall efficiency. Any seal leakage also reduces overall efficiency.
(3) Two sets of three concentric and coaxial cylinders/bores are extremely difficult to fabricate with tight tolerances. Also, the manufacturing of two sets of three concentric and coaxial pistons/rods to tight tolerances is quite difficult. Further, minimizing the stack-up of tolerances when the piston assembly must reciprocate within the nest of cylinders without binding on the one hand and without high leakage due to the large clearances on the other hand, is extremely challenging.
(4) The pumping pistons must be larger in diameter to maintain a needed piston pumping area than would be necessary without the connecting rod. The larger diameter pumping pistons produce higher friction and higher leakage. The diameter of the connecting rod must be relatively large since it must transmit the forces necessary to accelerate and decelerate the opposite side single free-piston assembly mass, which translates into an even larger increase in the pumping piston diameter.
(5) The structure of the assembly is not sufficiently rigid to allow acceptable ringless combustion, as will be further addressed later.
(6) The dual piston assembly is not mechanically balanced.
Accordingly, it is an objective of the present invention to provide for stoppage of a combustion piston and pumping piston in a free-piston engine at positions providing an appropriate top dead center position of the combustion piston.
Another objective of the present invention is to provide a free-piston engine which can be practically operated in a four-stroke cycle.
Yet another objective of the present invention is to provide a free-piston engine which is mechanically balanced.
Still another objective of the present invention is to provide a free-piston engine which is mass balanced.
Yet another objective of the present invention is to provide a free-piston engine which can be operated for a wide range of target compression ratios.
Still another objective of the present invention is to provide a free-piston engine assembly which is sufficiently rigid to allow for acceptable ringless combustion.
In order to achieve the foregoing objectives, in one aspect the present invention provides a free-piston engine including at least one dual piston assembly having a pair of axially opposed combustion cylinders and a free-floating combustion piston contained in each of the combustion cylinders for reciprocating linear motion responsive to combustion within the combustion cylinder. At least one pumping piston extends from and is fixed to each of the combustion pistons and each pumping piston is received within a hydraulic cylinder which is fixed in position between the paired combustion cylinders. A cage structure rigidly connects combustion pistons and surrounds the hydraulic cylinders and pumping pistons. As in conventional designs, ports in each of the hydraulic cylinders admit fluid at a first pressure and discharge fluid at a pressure higher than the inlet.
The hydraulic cylinders may be rigidly connected and the combustion pistons are rigidly connected by the cage structure so that when one of the combustion pistons is at top dead center, the other combustion piston is at bottom dead center.
The engine of the present invention may further include a bushing surrounding and guiding a rod interposed between and connecting a combustion piston with a pumping piston in order to allow for use of a ringless combustion piston.
The engine of the present invention is computer controlled with provision of position indicators on each cage connecting paired pistons, position sensors for reading the position indicators and an electronic control unit (ECU) for determining position of the cage, velocity, acceleration, et cetera and for controlling associated valving to stop movement of the dual piston assembly at TDC and BDC positions providing a target compression ratio.
In one preferred embodiment the engine of the present invention includes at least two of the dual piston assemblies and a synchronizer connecting the cages for synchronized parallel movement of the dual piston assemblies in opposite directions. The synchronizer can be the combination of a rack on each of the cages and a pinion located between and engaged by the racks, a chain/sprocket assembly or other similar means.
In another aspect, the present invention provides a method of operating a free-piston engine having at least one dual piston assembly as described above. The method involves drawing a fluid at low pressure through a low pressure fluid intake valve, into the hydraulic cylinders as the pumping pistons travel from BDC to TDC and discharging the fluid at a higher pressure, as the pumping pistons travel from TDC to BDC. Position indicators on the piston assembly are read to generate position signals and, on the basis of those position signals, the ECU determines a stoppage position for the dual piston assembly which provides a target compression ratio. The ECU generates a command signal for closing the low pressure fluid intake valve in the current cycle, to cause the dual piston assembly to stop at the determined stoppage position and to thereby achieve the target compression ratio in real time. The stoppage position is determined to allow the low pressure fluid intake valve to remain open through completion of filling fluid of a hydraulic cylinder and to close the low pressure fluid valve during discharge back to low pressure, generally of between 20% and 100% (idle) of the filled volume of the hydraulic cylinder, depending primarily on engine load and system high pressure. In determining the command signal for closing the intake valve, the ECU may also utilize signals representing the low (inlet) and high (outlet) pressures of one or more hydraulic cylinders. One approach to determination of a target position for closing the intake valve involves determination of energy produced by a single combustion event in a given cycle, as a function of velocity and acceleration of a dual piston assembly.
Preferably, the method of the present invention further includes a failsafe feature in which a range of closing positions for the low pressure fluid intake valve is determined on the basis of engine operating parameters such as fuel supply rate and the high (outlet) pressure of one or more hydraulic cylinders. In this preferred embodiment, the engine is shut off when the detected stoppage position is outside the established range for stoppage position.
The free-piston of the present invention further includes at least one fluid intake valve for controlling the emission of fluid into one of the hydraulic cylinders. In a preferred embodiment, that fluid intake valve is the fast acting valve disclosed in applicants"" prior U.S. Pat. No. 6,170,524, the teachings of which are incorporated herein by reference. In another preferred embodiment the fluid intake includes a valve member having a cupped head with a peripheral sealing surface and opposing concave and convex surfaces, and an integral guide stem extending from the convex surface. This preferred embodiment of the intake valve further includes a guide member with an axial bore receiving the guide stem of the valve member and providing for axial reciprocating movement of the guide member relative thereto between open and closed positions. A spring is included for biasing the valve member toward the closed position where the sealing surface of the head seals against a valve seat. The valve seat surrounds an axially extending port in fluid communication with one of the hydraulic cylinders. A reciprocal pin is mounted coaxially within the port for reciprocating movement between a retracted position and an extended position wherein the pin is in contact with the concave surface of the cupped head and holds the valve member in the open position. This preferred valve structure further includes an outlet port which may optionally be connected to a fluid accumulator which, in turn, may include a gas-filled bladder. A fluid connector connects TDC space within one cylinder with the axial bore of the guide member so that, as fluid pressure within the one cylinder is increased as the pumping piston therein approaches top dead center, the increased pressure operates on the guide stem to force the valve member into its closed position.
In another preferred embodiment, the free-piston engine of the present invention further includes impact pads mounted on the cage (5) for limiting movement of the dual piston assembly into the combustion cylinders.
Optionally, the dual piston assembly may further include balancing members mounted on opposing sides of and geared to the dual piston assembly for reciprocating motion in a direction opposite to the direction of motion of the dual piston assembly.
In yet another embodiment the free-piston engine of the present invention includes four parallel, side-by-side combustion cylinders, each having a free-floating combustion piston mounted therein for reciprocating linear motion, responsive to successive combustions within the combustion cylinders. As in the previously described embodiments, at least one pumping piston extends from and is fixed to each of the combustion pistons and a hydraulic cylinder is provided for receiving each of the pumping pistons. In this preferred embodiment a shuttle cylinder is axially aligned with and is in fluid communication with each of the hydraulic cylinders. A shuttle piston is mounted in each shuttle cylinder for reciprocating motion therein. Connectors rigidly and axially connect a shuttle piston to each of the pumping pistons. Transfer tubes provide fluid communication between first and second shuttle cylinders and between third and fourth shuttle cylinders. Flexible linkages are arranged within and run through the respective transfer tubes and are connected to the shuttle pistons of the first and second shuttle cylinders and the shuttle pistons of the third and fourth shuttle cylinders, respectively. A linkage connects the shuttle pistons in the second and third shuttle cylinders for movement together in tandem along with their associated pumping pistons and combustion pistons.
In still another preferred embodiment of the present invention, four of the dual piston assemblies are axially paired with one pair of dual piston assemblies in parallel with the other pair of dual piston assemblies. This embodiment further includes an outer cage rigidly affixed to one of the cages in the axially paired dual piston assemblies. A synchronizer, similar to that mentioned above, connects the two outer cages for synchronized movement in opposite directions. As is the case of the synchronizer described in connection with other embodiments, this synchronizer may include a rack on each of the outer cages and a pinion arranged between and engaged by each of the racks.